The present invention relates to compressor engineering, more particularly, to methods of regulating multistage axial compressor output and to axial compressors designed to carry into effect said methods.
This invention can be used to advantage in axial compressors designed for the following applications:
Compressing atmospheric air to be fed into oxygen plant separation units.
Compressing forced blast to be fed into blast furnaces and other metallurgical objects.
Compressing atmospheric air or air-oxygen mixture to be fed into magnetohydrodynamic generators.
Compressing air or another gas for use in chemical industries.
Compressing gas in other applications where large gas consumption has to be varied over a wide range.
The compressors used in the abovesaid applications must of necessity meet the requirement of providing a wide range of output regulation at a constant delivery pressure. The range of regulation is determined by the process requirements involved and by the need to compensate for seasonal changes in the temperature of the ambient air, amounting to 40-50 percent of the maximum output.
Axial compressors have high capacity and efficiency and, therefore, lend themselves to industrial applications, particularly with a view to the tendency of intensification of production processes and enlargement of the equipment employed. However, the use of axial compressors is hampered by their main disadvantage, viz., the comparatively narrow working range.
In multistage axial compressors the process of gas compression takes place in each stage successively. It is known that the performance of a compressor stage is determined by the flow coefficient which is essentially the ratio of the flow velocity axial component to the rotational speed of the rotor. When the flow coefficient decreases, the angle of incidence and the stage head increase until the angle of incidence reaches the critical value at which stall occurs. This phenomenon sets limitation on the minimum flow through the stage. When the flow coefficient increases, the stage head and the angle of incidence decrease until the stage becomes choked, i.e. the flow coefficient is at the maximum possible value and limits the flow through the stage.
The dimensions of each stage for the design conditions are suited to optimum distribution of the flow coefficients through the stages. Deviation from the design working conditions results in deviation from the desired flow coefficients or, in other words, in mismatching of the compressor stages.
For example, reducing the compressor output with the rotor speed constant causes decrease in the flow coefficient and increase in the first stage head which brings about increase in the pressure and unit weight of the gas at the inlet to the second stage. Consequently, the decrease in the flow coefficient of the second stage is greater than that of the first stage and so forth, the mismatching increasing from stage to stage. The greatest decrease of the flow coefficient is in the last stage where stall conditions cause surge and thereby restrict the reduction of the compressor output.
Increasing the compressor output causes increase in the flow coefficient, decrease in the first stage head and decrease in the pressure and unit weight of the gas at the inlet to the second stage, which causes still further increase in the flow coefficient of the second stage, and so forth. The greatest increase of the flow coefficient is in the last stage where choking conditions restrict the increase of the compressor output.
Mismatching of the compressor stages is also unavoidable if varying the rotor speed in proportion to the compressor output is resorted to in order to maintain the design flow coefficient of the first stage. For example, the rotor speed increases directly as the compressor output, whereas the first stage head, and the pressure and unit weight of the gas at the inlet to the second stage increase as the square of the speed. The flow coefficient of each subsequent stage decreases, the greatest decrease being in the last stage where stall conditions restrict the increase of the compressor output. If the compressor output and the rotor speed are decreased, the compression in the stages decreases and the flow coefficient increases through the stages. With the last stage choked, the coefficient of flow therethrough is at its maximum and cannot increase any more. Therefore, further reduction of the rotor speed causes the first stage flow coefficient to decrease until a rotating stall occurs in the first stage whereby the range of the rotor working speeds is restricted.
Thus, the increasing mismatching of the compressor stages due to the effect of gas compressibility in the process of compression restricts the range of the compressor output and the range of the rotor speed, the restriction being engendered on the one hand by choking conditions and on the other hand by stalling conditions in one of the compressor stages.
The effect of gas compressibility increases with the compression ratio (the ratio of the delivery pressure to the suction pressure). Therefore, the higher the delivery pressure, the narrower the working range of a multistage axial compressor.
The mismatching of the stages of a multistage axial compressor will be brought to a minimum, provided that the compressor output and the rotor speed vary in such a manner as to maintain the design flow coefficient of the middle stage. Under such conditions, the compressor performance curve showing the relation of the delivery pressure to the compressor output approximates to a parabolic line. When a compressor is used in a gas turbine installation, the performance curve showing the relation of the delivery pressure to the output is most nearly parabolic and it is this feature that renders axial compressors most suitable for use in gas turbine engines.
However, with a high compression ratio the working range of axial compressors becomes so narrow that the rotating stall occurring in the first stages of the compressor at low rotor speeds makes it impossible to start the engine without recourse to special means.
With compression ratios up to 8-12, recourse is made to blowing off air after the compressor intermediate stage in order to increase the flow through the first stages, decrease the associated angle of incidence and eliminate rotating stall. Alternatively, the pitch of the inlet guide vanes is altered so as to change the direction of the air flow and the angle of incidence on the first-stage vanes. With compression ratios above 8-12, these measures do not suffice and recourse is made to a two-spool compressor layout the principle of which consists in the following:
Inasmuch as the rotating stall in the first stage at a low rotor speed entails the choking of the last stage restricting the rate of flow through the compressor, it is desirable that during the starting of a gas turbine installation the first and last compressor stages run at different rotational speeds. Then, increasing the speed of the last stage rotor with unchanged output will decrease the flow coefficient of the last stage, eliminate choking and permit of further increasing the output. At the same time, decreasing the speed of the first stage rotor will increase the first stage flow coefficient, thereby diminishing the angle of incidence on the first stage and eliminating the rotating stall.
A method of regulating the output of a multistage, two-spool, axial compressor with two telescopically mounted turbine shafts is known. Each of the turbines runs at a variable speed and drives the respective compressor spool. Regulation is effected by varying the rotational speeds of the first and second spools, the speed increment of the second spool being greater than that of the first spool in virtue of the appropriate turbine power balance. This regulation decreases mismatching of the stages of the first and second spools during starting and at off-design operating conditions. In connection with a multistage axial compressor used in a gas turbine installation, such regulation decreased mismatching of the compressor stages and wides the range of working conditions.
The advanced variation of the second spool rotational speed allows of increasing the flow through the stages of the first spool and the low coefficient of the first stages, which provides design angles of incidence and shifts the limit of rotating stall into the area of lower rotational speeds. Yet, the widening of the working range of a two-spool compressor does not result in obtaining a wide range of output variation because the latter remains substantially narrow with a high compression ratio.
The narrow working range of multistage axial compressors is a disadvantage in connection with gas turbine installations and a still greater disadvantage as regards the employment of said compressors in other industrial applications. Therefore, it is an urgent problem today to provide for varying compressor output over a wide range, particularly at the required constant delivery pressure.
The generally known method of regulation by varying the speed of the rotor, or two rotors in a two-spool layout, gives only a limited range of output variation. For example, in the case of axial compressors with a moderate compression ratio of 3-4.5 the range of output variation at a constant delivery pressure is only 20-25 percent of the maximum output, this figure decreasing with increasing compression ratio.
Moreover, this method suffers from such disadvantages as drive losses and difficulty of turning the vibration of the compressor vanes in a wide range of rotational speeds.
For this reason the development of the methods of regulating the output of the axial compressors employed in the industry has tended towards wide-range regulation of compressor output at a constant rotor speed. With a constant-speed rotor running, drive can be effected by the use of an electric motor, which is conductive to operating economies.
Also known in the art is a method of regulating the output of a constant-speed axial compressor by bleeding some of the delivery to a recuperative turbine mounted on the shaft of said compressor, which recuperative turbine returns to the compressor shaft the energy expended in compressing the bled delivery, thereby rendering the regulation of the compressor output more economical. However, this method has not been accepted widely, since from the economy viewpoint, it is inferior to the method of regulating the output of an axial compressor by changing the pitch of the stator vanes.
The method of regulating the output of a multistage axial compressor at a constant rotational speed by changing the pitch of the stator vanes is most economical and is used nowadays by all the leading turbine manufacturers.
There is known a method of regulating the output of an industrial multistage axial compressor by changing the pitch of the stator vanes.
With this method, the pitch of the stator vanes is changed so as to provide optimum angles of incidence or attack in the compressor stages in off-design operating conditions. This expedient reduces or fully eliminates mismatching of the compressor stages in varying the compressor output and provides a wide range of regulating the compressor output. Changing the pitch of the stator vanes alters the angular direction of the issuing flow, i.e. alters the peripheral speed component of the flow at the inlet to the next rotor wheel, whereby variation is caused in the angle of incidence or attack and in the head of the compressor stage being regulated. The angles of the vanes generally differ between the stages of a multistage compressor because of different amounts of mismatching. The number of regulated stages increases with the range of output regulation, covering, as a rule, 40 to 100 percent of the stages on industrial axial compressors.
Changing the pitch of only the inlet guide vanes, frequently resorted to in the axial compressors of gas turbine engines to guard against rotating stall in starting the engine, or prime moved provides an insufficient range of regulation of the compressor output at a constant rotor speed and, therefore, this method is not used in connection with industrial axial compressors.
In a multistage axial compressor comprising variable-pitch stator vanes, the vanes are mounted in a bearing supported by a bearing support installed inside the compressor casing and forming the periphery of the blading. To the root end of each vane is rigidly attached a pitch control lever the other end of which has a pin fitting into a circular groove in a cylindrical sleeve and adapted to move in said groove during the movement of said cylindrical sleeve. The cylindrical sleeve is located inside the compressor casing between the bearing support and the inside of the casing and is kinematically connected to a servomotor designed to move it axially. The movement imparted to the cylindrical sleeve causes the pitch control lever pins to move and thereby turn the vanes about their axes, changing the vane pitch.
The method of regulating the output of a multistage axial compressor at a constant rotor speed by changing the pitch of the stator vanes meets the working requirements of industrial compressors, but it complicates compressor construction and causes increase in manufacturing labour and materials. Furthermore, cost is raised by the employment of the vane pitch change mechanism which comprises a large number of precision elements which have to be carefully manufactured, assembled and adjusted. A still further disadvantage is poor operating reliability: wear and seizure of even one variable-pitch vane will necessitate compressor repairs.
Endeavour to carry into effect the same principle of regulation by simpler constructional means has brought about a method of regulating compressor output by rotating the inlet guide vanes. By this method, changing the direction of the flow or the peripheral component of the flow velocity is obtained by imparting the flow the transfer speed of the rotation of the inlet guide vanes, the speed of vane rotation being chosen so as to provide the optimum angle of incidence of the rotor wheel blades. The arrangement for carrying this method into effect comprises inlet guide vanes positioned before the rotor wheel and mounted on a sleeve installed on a shaft coaxial with the wheel shaft. The inlet guide vane assembly is driven, for example, by a controllable hydraulic motor which rotates the inlet guide vane assembly at different speeds in the forward and reverse directions relative to the wheel shaft. This produces variation in the peripheral component of the flow velocity and in the angular direction of the flow at the inlet to the wheel, whereby the operation of the compressor is regulated.
The effect of varying the peripheral component of the flow velocity and the angular direction of the flow downstream of a vane row rotating at variable speed has found use for regulating fans comprising two coaxial impellers rotating in opposite directions, the regulation being effected by changing the rotational speed of at least one of the impellers.
However, the employment of rotating guide vanes in multistage axial compressors makes the compressor construction still more complicated than in the case of the variable pitch vanes, there being a rotor and a stator rotating simultaneously. The need for each of the guide vane assemblies to run generally at its own speed complicates the compressor construction still further. The construction of a multistage axial compressor can be simplified by using one rotating inlet guide vane row, but, as stated above, this means is not sufficient to obtain a wide range of output regulation.
Thus, in the prior art, the economical methods of widerange regulation of the output of multistage axial compressors entail substantial complication of the compressor construction.